This invention relates to a twin screw fluid machine and, more particularly, to a screw fluid machine having a casing with a bore shape suitable for the realization of high efficiency.
Details of a fundamental structure of a screw fluid machine are disclosed in U.S. Pat. No. 3,423,017. Generally, in a fluid machine which pumps gas such as a compressor, expander, or vacuum pump, the temperature of the gas is high on the high-pressure side. For example, in the case of a nonlubricated screw compressor which compresses air at a compression ratio of 8, there is a possibility of the temperature of air on the highpressure side exceeding 300.degree. C. during operation, thereby increasing the thermal expansion of the rotor. A type of machine, such as that disclosed in U.S. Pat. No. 4,475,878, is known, in which the rotor is tapered so that its outside diameter at the high pressure side is reduced relative to that at the low-pressure side. According to this related art, the inside diameter of the bore which accommodates the rotor is increased by preliminarily calculating the thermal expansion of the rotor which occurs during the operation of the machine.
In traditional theory, it is thought that the thermal deformation of the casing which accommodates the rotor is small because the casing is cooled by a water jacket and by radiation from the surface of the casing. However, it has been found by measuring the temperature distribution over the casing using sensors embedded in different portions of the casing that the temperatures of these portions greatly differ from each other.
FIG. 15 shows the distribution of temperatures of a casing along a cross section of the bore perpendicular to the axis thereof. In FIG. 15 are illustrated bore walls 81 and 82 which face male and female rotors respectively, theoretical axes 83 and 84 of the male and female rotors, a high-pressure-side fluid passage 85 (hereinafter referred to as "high-pressure port") formed at an intersection of the bore walls 81 and 82, and a peripheral direction distribution curve 86 of the temperature of the bore wall 81 on the side of the male rotor. This curve represents a bore wall temperature T at a point A by the length of a line segment AB defined on a straight line which passes through an axis 83 of the male rotor and the point A. As is understood from FIG. 15, the bore wall temperature T on the side of the male rotor is high in the vicinity of the high-pressure port 85 and becomes lower as an angle .theta. in this figure decreases.
FIG. 16 shows the distribution of the temperature of the bore wall in the direction of the axis of the bore in a plane which contains the axis of the rotor. In FIG. 16 are illustrated a low-pressure end surface 88 of the bore wall 81, a high-pressure end surface 89 of the bore wall 81, an axis 90 of a male rotor 91, and a straight line 92 which represents the distribution of the temperature of the bore wall 81 in the direction of the axis of the bore and which represents a bore wall temperature T at a point D by the length of a line segment DE defined on a straight line which is perpendicular to the axis 90. As can be understood from FIG. 16, the bore wall temperature T is high at the high-pressure side and is low at the low-pressure side.
The conventional bore wall 81 is formed in such a manner that the inside diameter is uniform relative to the axis 90 so as to form a truly round cross section of the bore. However, the bore wall 81 is deformed by a change in the temperature thereof during operation, as mentioned above, so that the shape of the bore 81 deviates from the round even in a plane perpendicular to the center axis of the bore.
This related art has been described with respect to the male rotor alone, but it goes without saying that these facts also apply with respect to the portion on the side of the female rotor.
FIG. 17 shows gaps between the bore wall and rotors in accordance With the related art. In FIG. 17 are illustrated lines 93 and 94 which indicate the outside diameters of the male and female rotors at an ordinary temperature, lines 95 and 96 which indicate the outside diameters of the male and female rotors when the rotors are thermally deformed during operation, lines 98 and 99 which indicate the inside diameters of the bore walls on the sides of the male and female rotors at the ordinary temperature, and lines 100 and 101 which indicate the inside diameters of the bore walls on the sides of the male and female rotors when the bore walls are thermally deformed during operation The inside-diameter lines 98 and 99 of the bores on the sides of the male and female rotors are circular at the ordinary temperature.
During operation, the bore walls 81 and 82 are thermally deformed in accordance with the temperature distribution indicated in FIG. 16, and each point on the inside-diameter lines of the bore walls 95 and 96 is displaced outward in the radial direction. Specifically, this displacement is large in the vicinity of the highpressure outlet 85. Since the rotor is a revolving body, the outside-diameter line of the rotor when thermally deformed forms a circle in a plane perpendicular to the axis thereof. Accordingly, as shown in FIG. 17, a gap h between the outside-diameter lines 95 and 96 of the male and female rotors and the inside-diameter lines 100 and 101 of the bore walls is maximum in the vicinity of the highpressure port 85.
The casing in accordance with the related art is formed in such a manner that the inside diameters of the bore walls 81 and 82, which form cylindrical surfaces, are increased by preliminarily calculating the thermal expansions of the rotors exhibited during operation. However, as described above, the inside-diameter lines of the bore walls 100 and 101 are not deformed uniformly, and, therefore, the gap h shown in FIG. 17 becomes nonuniform. A leak which occurs at the gap h flows from a groove of one of the rotors to the adjacent groove over the top of a lobe formed therebetween (refer to FIG. 1). If, as described above, the gap h is large on the highpressure side, the loss of power is very large because the difference between the pressures in adjacent grooves is large on the high-pressure side. That is, a large difference between the ends of the course of the leak causes a large amount of leak per unit time and, hence, a large energy loss due to the leak.